Compression ignition engine

ABSTRACT

In a compression ignition engine employing a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of an intake valve lift and an intake valve closure timing, a control system operates to temporarily lower an effective compression ratio of the engine by controlling the intake valve characteristic during a cranking period of cold starting operation. At a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise, the effective compression ratio is risen by controlling the intake valve characteristic. After combustion of the engine has been stabilized, the intake valve characteristic is brought closer to a desired value determined based on engine operating conditions by way of closed-loop control.

TECHNICAL FIELD

The present invention relates to a compression ignition engine employing a variable valve operating system for at least one of intake and exhaust valves, and specifically to the improvement of a compression ignition engine control technology suited to compression ignition engines such as a four-stroke-cycle Diesel engine, a two-stroke-cycle Diesel engine, a premix compression ignition engine, and the like.

BACKGROUND ART

In recent years, there have been proposed and developed various engine control technologies for compression ignition engines with variable valve operating systems. Generally, a variable valve operating system, capable of variably adjusting a valve lift and valve timing of at least one of intake and exhaust valves of a reciprocating internal combustion engine depending on engine operating conditions, is widely utilized for controlling a charging efficiency, an effective compression ratio, and an amount of residual gas of the engine, thereby enhancing the engine power performance and exhaust emission control performance. In Diesel engines or premix compression ignition engines, air alone is compressed during the compression stroke, and then fuel, which is sprayed or injected into the cylinder, is self-ignited due to a temperature rise of the compressed air (heat produced by compressing the incoming air). That is, such self-ignition of the sprayed fuel can be performed under a high-temperature high-pressure condition where the pressure and temperature of the compressed air are high enough to ignite spontaneously the sprayed fuel. The spontaneous ignition temperature and spontaneous ignition pressure needed for self-ignition both change depending on sorts of fuel that is sprayed into the compressed air. Generally, unless the temperature of the compressed air is more than 1000 degrees K (Kelvin temperature) and the pressure of the compressed air is more than 1 MPa (mega Pascal), it does not result in spontaneous ignition of the sprayed fuel.

For the reasons discussed above, the compression ratio of the engine has to be set to a high ratio of 15:1 or more, so that the in-cylinder pressure and in-cylinder temperature become high enough to spontaneously ignite the sprayed fuel and to achieve the combustion of the sprayed fuel, for instance, even when the engine cylinder wall temperature is still low during cold starting and thus heat of the compressed air is taken by the cylinder wall. However, such a high compression ratio causes excessively high pressures acting on the piston after the engine warm-up has been completed, thus resulting in the increased mechanical friction loss and reduced engine power performance. To avoid this (for avoidance of undesirable mechanical friction loss), it is effective to reduce the compression ratio to 15:1 or less after completion of the engine warm-up, in other words, after the engine starting operation has been completed, thereby enhancing the engine performance. After completion of the starting operation, the cylinder wall temperature becomes high, and thus the heat produced by compressing the air is hard to be taken by the cylinder wall even at a comparatively low compression ratio. As a result, the temperature and pressure of the compressed air easily become high during the compression stroke, thus ensuring self-ignition of the sprayed fuel. As is generally known, the variable compression ratio adjustment can be achieved by mechanically varying the clearance volume, that is, the air volume with the piston at top dead center (TDC). Alternatively, the variable compression ratio adjustment can be achieved by mechanically varying the piston stroke characteristic. However, such variable compression ratio devices, for example, a multi-link variable compression ratio device and the like, capable of mechanically varying the clearance volume or mechanically varying the piston stroke characteristic, have complicated mechanical configuration and structure. In lieu thereof, it is possible to variably adjust the mass of air entering the engine cylinder at the beginning of the compression stroke by retarding or advancing the intake-valve closure timing, denoted by “IVC” and expressed in terms of crank angle. In such a case, it is possible to retard a rise in in-cylinder pressure and a rise in in-cylinder temperature with respect to a predetermined crank angle. In other words, it is possible to lower the effective compression ratio by retarding an in-cylinder pressure rise and/or an in-cylinder temperature rise by way of variable adjustment of intake valve closure timing IVC. One such IVC adjustment type variable compression ratio device for a compression ignition engine has been disclosed in Japanese Patent Provisional Publication No. 1-315631 (hereinafter is referred to as “JP1-315631”). In the case of JP1-315631, the IVC adjustment type variable compression ratio device is exemplified in a two-stroke-cycle Diesel engine. Concretely, when it is determined that the current operating condition of the two-stroke-cycle Diesel engine corresponds to an engine starting period, intake valve closure timing IVC is phase-advanced towards a timing value near bottom dead center (BDC) by means of an electric-motor driven variable valve operating device (or a motor-driven variable valve timing control (VTC) system), thereby increasing an effective compression ratio and consequently enhancing the self-ignitability during the starting period. In contrast, during engine normal operation, intake valve closure timing IVC is phase-retarded to decrease the effective compression ratio and consequently to reduce a fuel consumption rate. The motor-driven VTC system of JP1-315631 uses a rotary-to-linear motion converter, such as a ball-bearing screw mechanism, for changing relative phase of an intake-valve camshaft to an engine crankshaft. The rotary-to-linear motion converter (the ball-bearing screw mechanism) of JP1-315631 is comprised of a warm shaft (i.e., a ball bearing shaft with helical grooves) driven by a step motor, an inner slider (i.e., a recirculating ball nut), recirculating balls provided in the helical grooves, and an outer slider axially movable together with the inner slider and rotatable relative to the inner slider. The other types of variable valve operating devices have been disclosed in (i) Japanese document “JSAE Journal Vol. 59, No. 2, 2005” published by Society of Automotive Engineers of Japan, Inc. and titled “Gasoline Engine: Recent Trends in Variable Valve Actuation Technologies to Reduce the Emission and Improve the Fuel Economy” and written by two authors Yuuzou Akasaka and Hajime Miura, and (ii) Japanese document “Proceedings JSAE 9833467, May, 1998” published by Society of Automotive Engineers of Japan, Inc. and titled “Reduction of the engine starting vibration for the Parallel Hybrid System” and written by four authors Hiroshi Kanai, Katsuhiko Hirose, Tatehito Ueda, and Katsuhiko Yamaguchi. The Japanese document “JSAE Journal Vol. 59, No. 2, 2005” discloses various types of variable valve operating systems, such as a helical gear piston type two-stepped phase control system, a rotary vane type continuously variable valve timing control (VTC) system, a swing-arm type stepped valve lift and working angle variator, a continuously variable valve event and lift (VEL) control system, and the like. The VTC and VEL control systems are operated by means of respective actuators for example electric motors or electromagnets, each of which is directly driven in response to a control signal (a drive signal) from an electronic control unit (ECU). Alternatively, the VTC and VEL control systems are often operated indirectly by means of a hydraulically-operated device, which is controllable electronically or electromagnetically. On the other hand, the Japanese document “Proceedings JSAE 9833467, May, 1998” teaches the use of a variable valve timing control system installed on the intake valve side of an engine of a hybrid vehicle employing a parallel hybrid system, for prevention of rapid engine torque fluctuations, which may occur during engine stop and start operation.

SUMMARY OF THE INVENTION

In the case of the compression ignition engine with the variable valve operating device, as disclosed in JP1-315631, the effective compression ratio is controlled to a relatively high ratio by means of the variable valve operating device during the engine starting period. After the starting operation has been completed, the effective compression ratio is controlled to a relatively low ratio by means of the variable valve operating device. In such an engine control system, there is an increased tendency for the work of compression to increase during the engine starting period. The increased work of compression leads to a drop in cranking speed, thereby resulting in an increased heat loss of the compressed air (compressed gas). As a result of this, a compression temperature, i.e., a temperature of the compressed gas, tends to drop, thus deteriorating the engine startability. According to the engine control system as disclosed in JP1-315631, the effective compression ratio is lowered and decreasingly compensated for, just after the starting operation has been completed. That is to say, the effective compression ratio is controlled to a relatively low ratio, though there is a possibility that the combustion stability is still insufficient just after completion of the starting operation. This leads to the problem of deteriorated combustion stability. Additionally, in order to increase cranking speed, the compression ignition engine as disclosed in JP1-315631 often uses an engine starter of a high torque capacity (a motor generator of a high torque capacity in case of a hybrid vehicle). This leads to another problem of increased manufacturing costs and increased weight. Instead of using an engine starter of a high torque capacity, a so-called decompression device can be used to increase cranking speed. The decompression device is often used for an engine for a two-wheeled vehicle, so as to constantly open an exhaust valve during cranking, thus reducing the work of compression and consequently increasing the cranking speed. However, the decompression device itself does not have an effective-compression-ratio reducing function that reduces the effective compression ratio after completion of the starting operation. Thus, it is difficult to realize the improved fuel economy (i.e., the reduced fuel consumption rate) during normal engine operation, by the use of the decompression device.

In more detail, in the VTC system disclosed in JP1-315631, when there is no application of electric current to the step motor of the VTC system and thus the step motor is de-energized (OFF), intake valve closure timing IVC is automatically controlled to a timing value near bottom dead center (BDC), for example, 20 degrees of crank angle after BDC, under an unfailed condition of the VTC system. Conversely when the step motor is energized (ON), intake valve closure timing IVC is controlled to a timing value retarded from the piston BDC position, for example, 60 degrees of crank angle after BDC. JP1-315631 teaches the phase-advance of intake valve closure timing IVC to a timing value near BDC during the engine starting period, and also teaches the phase-retard of intake valve closure timing IVC after completion of the starting operation. However, according to the system of JP1-315631, the effective compression ratio remains high during cranking, thus resulting in an undesirable drop in cranking speed.

In the case of the system as disclosed in the Japanese document “JSAE Journal Vol. 59, No. 2, 2005”, intake valve closure timing IVC is not phase-retarded from BDC during cranking and cold starting with a starter energized. The effective compression ratio remains high during the cranking and starting period. This also leads to the problem of reduced cranking speed.

In the case of the system as disclosed in the Japanese document “Proceedings JSAE 9833467, May, 1998”, intake valve closure timing IVC of the starting period is phase-retarded to reduce the quantity of air charged in the engine, thus preventing a rapid rise in torque generated by the engine. However, even after cranking operation, intake valve closure timing IVC remains retarded, thus deteriorating the engine startability or self-ignitability during start operation.

It is, therefore, in view of the previously-described disadvantages of the prior art, an object of the invention to provide a compression ignition engine, capable of avoiding the aforementioned problem that spontaneous ignition of fuel does not take place owing to a drop in cranking speed during a starting period.

In order to accomplish the aforementioned and other objects of the present invention, a compression ignition engine comprises sensors that detect engine operating conditions, a variable valve operating system comprising at least a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and actuated by an actuator, and a control unit configured to be electrically connected to the sensors and the actuator for controlling the variable valve actuation mechanism via the actuator to bring the intake valve characteristic closer to a desired value determined based on the engine operating conditions detected by the sensors, the control unit comprising a processor programmed to perform the following, temporarily lowering an effective compression ratio of the engine by controlling the intake valve characteristic during a cranking period of cold starting operation, rising the effective compression ratio by controlling the intake valve characteristic at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise, and bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.

According to another aspect of the invention, a compression ignition engine comprises sensors that detect engine operating conditions, a variable valve operating system comprising at least a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and actuated by an actuator, a decompression device provided to operate an exhaust valve in a decompression mode corresponding to a constantly-opened valve operating state during a cranking period of cold starting operation, and a control unit configured to be electrically connected to the sensors and the actuator for controlling the variable valve actuation mechanism via the actuator to bring the intake valve characteristic closer to a desired value determined based on the engine operating conditions detected by the sensors, the control unit also configured to be electrically connected to the decompression device for switching the exhaust valve to the decompression mode during the cranking period, and the control unit comprising a processor programmed to perform the following, temporarily lowering an effective compression ratio of the engine by maintaining the exhaust valve in the decompression mode corresponding to the constantly-opened valve operating state during the cranking period, inhibiting the decompression mode and returning the exhaust valve to a normal operating state at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise, rising the effective compression ratio by controlling the intake valve characteristic substantially at the point of time when the predetermined cranking speed threshold value has been reached owing to the cranking speed rise, and bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.

According to a further aspect of the invention, a method for controlling a compression ignition engine employing a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve, the method comprises temporarily lowering an effective compression ratio of the engine by controlling the intake valve characteristic during a cranking period of cold starting operation, rising the effective compression ratio by controlling the intake valve characteristic at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise, and bringing the intake valve characteristic closer to a desired value determined based on engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.

According to a still further aspect of the invention, a method for controlling a compression ignition engine employing a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and a decompression device provided to operate an exhaust valve in a decompression mode corresponding to a constantly-opened valve operating state during a cranking period of cold starting operation, the method comprises temporarily lowering an effective compression ratio of the engine by maintaining the exhaust valve in the decompression mode corresponding to the constantly-opened valve operating state during the cranking period, inhibiting the decompression mode and returning the exhaust valve to a normal operating state at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise, rising the effective compression ratio by controlling the intake valve characteristic substantially at the point of time when the predetermined cranking speed threshold value has been reached owing to the cranking speed rise, and bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.

The other objects and features of this invention will become understood from the following description with reference to the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a system block diagram illustrating an embodiment of a compression ignition engine.

FIG. 2 is a construction drawing showing an electric-motor driven variable valve operating system, which is applicable to the compression ignition engine of the embodiment.

FIG. 3 is a construction drawing showing another electric-motor driven variable valve operating system, which is applicable to the compression ignition engine of the embodiment.

FIG. 4 is a construction drawing showing a camshaft sensor incorporated in the engine control system of the compression ignition engine of the embodiment.

FIG. 5 is a construction drawing showing another camshaft sensor incorporated in the engine control system of the compression ignition engine of the embodiment.

FIG. 6 is a crank-angle versus camshaft sensor signal characteristic diagram.

FIG. 7 is a diagram of intake valve open timing (IVO), intake valve closure timing (IVC), exhaust valve opening timing (EVO), and exhaust valve closure timing (EVC) in a four-stroke-cycle compression ignition engine.

FIG. 8 is a diagram of intake valve open timing (IVO), intake valve closure timing (IVC), exhaust valve opening timing (EVO), and exhaust valve closure timing (EVC) in a two-stroke-cycle compression ignition engine.

FIG. 9 is a time chart showing one phase-control characteristic of a variable valve operating system (a VTC system) incorporated in the engine control system of the compression ignition engine of the embodiment.

FIG. 10 is a phase-control characteristic diagram showing phase changes attained by the intake-valve VTC system, for effective compression ratio changes under various operating conditions, such as during cranking, after engine warm-up, and at maximum phase-advance timing.

FIG. 11 is a flow chart showing a starting-period VTC control routine executed within an electronic control unit incorporated in the engine control system of the compression ignition engine of the embodiment.

FIG. 12 is a disassembled view showing the detailed structure of a hydraulically-operated rotary vane type VTC mechanism, which is applicable to the engine control system of the compression ignition engine of the embodiment.

FIGS. 13A-13C are explanatory views showing the operation of a hydraulic control system for the hydraulically-operated rotary vane type VTC mechanism shown in FIG. 12.

FIG. 14A is an explanatory view showing the operating angular range of the rotary vane of the hydraulically-operated VTC mechanism shown in FIG. 12.

FIGS. 14B-14C respectively show the maximum phase-advance position and the maximum phase-retard position with regard to the rotary vane of the hydraulically-operated VTC mechanism shown in FIG. 12.

FIG. 15 is a time chart showing another phase-control characteristic of the VTC system.

FIG. 16 is a time chart showing another phase-control characteristic in the case of a combination of the VTC system and a decompression device.

FIG. 17 is a time chart showing another phase-control characteristic of the VTC system.

FIG. 18 is a valve event and lift control characteristic diagram showing valve event and lift characteristics attained by a continuously variable valve event and lift control (VEL) system, for effective compression ratio changes under various operating conditions, such as during cranking, after engine warm-up, and at maximum phase-advance timing.

FIG. 19 is a flow chart showing a glow-plug/electric-heater control routine.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring now to the drawings, particularly to FIG. 1, the variable valve operating system incorporated in the compression ignition engine of the embodiment is exemplified in a four-stroke-cycle engine. As indicated by the arrow in the system block diagram of FIG. 1, a crankshaft 2 of an engine 1 rotates clockwise. As is generally known, a piston position at which a piston 3 has moved to the bottom of the cylinder of engine 1, corresponds to 180 degrees of crank angle. The lowest piston position is called “bottom dead center (BDC)”. A piston position obtained when engine crankshaft 2 further rotates and thus piston 3 has reached the top of the engine cylinder, corresponds to 360 degrees of crank angle. The highest piston position is called “top dead center (TDC)”.

In the case of usual diesel combustion, diesel fuel (fuel oil) is sprayed or injected via a fuel injection valve 4 into the cylinder during the compression stroke. Then, the sprayed fuel is self-ignited and combusted due to the high-temperature high-pressure compressed gas (heat produced by compressing the incoming air). On the other hand, in the case of premix compression ignition, fuel is sprayed or injected via fuel injection valve 4 into the cylinder during the intake stroke so that the sprayed fuel is sufficiently premixed with air charged in the cylinder. Residual gas is set to a comparatively large amount for a temperature rise in air-fuel mixture. When piston 3 moves up, a temperature rise and a pressure rise in premixed air-fuel mixture occur, thereby resulting in spontaneous ignition of the air-fuel mixture so that the mixture is combusted. A fuel injection amount and injection timing of fuel injection valve 4 included in the electronic injection control system are both controlled, responsively to a sensor signal from a crank angle sensor 5, by means of an electronic control unit (ECU) 6. The purpose of crank angle sensor 5 is to inform the ECU 6 of engine speed Ne as well as the relative position of crankshaft 2.

During start operation, an engine starter 7 is operated to crank the engine 1 or to turn the crankshaft 2. In the case of a hybrid-vehicle engine, rather than using starter 7, engine 1 is rotated by means of a motor generator. Additionally, during the starting period, an electric current is applied to a glow plug 8 for a temperature rise in glow plug 8 and for promotion of vaporization of fuel, thus supporting or assisting spontaneous ignition. Harmful exhaust emission gases such as carbon monoxide (CO), hydrocarbons (HCs), soot (particulate matter), nitrogen oxides (NOx), and the like, are filtered out and purified by means of a catalytic converter 301.

An intake valve 9 and an exhaust valve 10 are installed in the upper part of engine 1. Intake valve 9 is driven by an intake cam 11, whereas exhaust valve 10 is driven by an exhaust cam 12. Intake cam 11 is mechanically linked via a variable valve actuation mechanism (or a variable valve characteristic adjustment mechanism) 13 to a camshaft timing pulley 14. In the embodiment shown in FIG. 1, a hydraulically-operated rotary vane type variable valve timing control (VTC) mechanism is used as variable valve actuation mechanism 13. In lieu thereof, a variable valve lift (VVL) mechanism or a continuously variable valve event and lift (VEL) control mechanism may be used as variable valve actuation mechanism 13. Rotation of crankshaft 2 is transmitted via a timing belt, a timing chain or the like to camshaft timing pulley 14. In the shown embodiment, exhaust cam 12 is linked directly to camshaft timing pulley 14. Alternatively, exhaust cam 12 may be linked to camshaft timing pulley 14 through the VTC mechanism for intake cam 11. In lieu thereof, exhaust cam 12 may be linked to camshaft timing pulley 14 through a separate VTC mechanism differing from the VTC mechanism for intake cam 11.

A sensor signal from an engine temperature sensor (a water temperature sensor or an engine coolant temperature sensor) 15, which detects engine temperature Te, is input into ECU 6. A sensor signal from a camshaft sensor 16 of the VTC system is also input into ECU 6. Camshaft sensor 16 is located near the intake camshaft associated with intake cam 11. Camshaft timing pulley 14 is driven by the engine crankshaft at ½ the revolution speed of crankshaft 2. In the variable valve operating system of FIG. 1, intake cam 11 is rotated with a phase difference between an angular phase detected by crank angle sensor 5 and an angular phase detected by camshaft sensor 16. The valve-opening action of intake valve 9 is performed once for each two revolutions of crankshaft 2, for entry of air into the cylinder.

During rotation of camshaft timing pulley 14, exhaust cam 12 linked to camshaft timing pulley 14 is also driven. The valve-opening action of exhaust valve 12 is performed once for each two revolutions of crankshaft 2, for exhausting burned gas from the engine cylinder. As can be seen from the left-hand side of FIG. 1, an air flow sensor 17, a turbo charger 18, and an exhaust gas recirculation (EGR) valve 19 are installed in an induction system 20 and arranged upstream of intake valve 9. Air flow sensor 17 is provided for measuring the quantity Qa of air entering the engine cylinder. Additionally, as input information indicative of engine load, the input interface of ECU 6 receives a sensor signal from an accelerator position sensor 100 that detects an amount APS of depression of an accelerator pedal.

Variable valve actuation mechanism 13 (or the hydraulically-operated rotary vane type VTC mechanism in the engine control system of the compression ignition engine of the embodiment shown in FIG. 1) is a variable phase control means, which is operable simultaneously with cranking operation of engine 1. In the case of the hydraulically-operated VTC mechanism of the compression ignition engine of the embodiment shown in FIG. 1, the VTC mechanism is operated by hydraulic pressure produced by an oil pump of engine 1, and therefore the hydraulic pressure produced by the engine oil pump tends to drop during cranking operation. Due to such a drop in the supplied hydraulic pressure, the VTC system has uncertainty in controlling the valve timing (IVC and/or IVO) of intake valve 9. Under a particular condition where the VTC system has uncertainty in controlling the valve timing due to a drop in hydraulic pressure produced by the engine oil pump, for example, during cranking, a separate electric-motor driven hydraulic oil pump 302 is driven simultaneously with the ignition-switch turn-ON operation so as to quickly satisfactorily feed or deliver hydraulic pressure to the VTC mechanism.

As shown in FIG. 2, a relative phase change of a camshaft 310 to camshaft timing pulley 14, that is, a valve timing change of intake valve 9, may be achieved by using a motor-driven spiral disk type VTC mechanism, rather than using the hydraulically-operated rotary vane type VTC mechanism, whose detailed construction will be described later in reference to the disassembled view of FIG. 12. Actually, in the case of the motor-driven spiral disk type VTC mechanism of FIG. 2, the phase difference between camshaft 310 and camshaft timing pulley 14 can be varied by means of a linkage (a motion converter) 312. The radial outside portion of linkage 312 is mechanically linked to both of camshaft timing pulley 14 and a spiral disk 311, such that the radial outside portion of linkage 312 slides along a guide groove 313 formed in camshaft timing pulley 14 and also slides along a guide groove 314 formed in spiral disk 311. On the other hand, the radial inside portion of linkage 312 is fixedly connected to camshaft 310. When the phase angle of spiral disk 311 relative to camshaft timing pulley 14 varies, the radial position of the outside portion of linkage 312 with respect to the axis of camshaft 310 varies, and thus a phase change of camshaft 310 relative to camshaft timing pulley 14 occurs. There are various methods to vary the phase angle of spiral disk 311 relative to camshaft 310. In the case of the motor-driven spiral disk type VTC mechanism shown in FIG. 2, a hysteresis motor 315 is used as an actuator (a driving power source or an electrically-controlled actuator means). Hysteresis motor 315 can apply torque to a hysteresis member 316 in a spaced, contact-free relationship with hysteresis member 316, for varying the phase angle of spiral disk 311 relative to camshaft timing pulley 14. Assuming that the car battery voltage is higher than a specified voltage value, the motor-driven VTC mechanism can be certainly operated by means of hysteresis motor 315 from the time when engine 1 is cranked. As is generally known, the magnitude of torque acting on hysteresis member 316 increases, as the applied electric current to hysteresis motor 311 increases. The increased torque acts to rotate hysteresis member 316 against the spring force of a biasing means (a return spring). As a result, it is possible to continuously vary the relative phase of camshaft 310 to camshaft timing pulley 14 responsively to the current value of the applied current to hysteresis motor 311. Therefore, it is possible to accurately control or adjust the actual relative phase of camshaft 310 to camshaft timing pulley 14 to a desired value by controlling the applied current by way of closed-loop control (feedback control) in response to the sensor signal from camshaft sensor 16.

Referring now to FIG. 3, there is shown a modification of the motor-driven VTC mechanism, which is applicable to the compression ignition engine of the embodiment. In the modified VTC mechanism of FIG. 3, relative phase of camshaft 310 to camshaft timing pulley 14, in other words, relative phase of camshaft 310 to crankshaft 2 is varied by means of a helical spline mechanism 320. Helical spline mechanism 320 is comprised of a substantially ring-shaped axially-movable helical-gear nut having an internal helically-splined groove portion, and an external helically-splined shaft end portion of camshaft 310. The internal helically-splined groove portion of the nut is in meshed-engagement with the external helically-splined shaft end portion of camshaft 310. Axially leftward movement or axially rightward movement of the nut of helical spline mechanism 320 causes a change in relative phase of camshaft 310 to camshaft timing pulley 14. As an actuator (a driving power source or an electrically-controlled actuator means) that creates axial movement of the nut of helical spline mechanism 320, a reversible motor 321 is used. As clearly shown in FIG. 3, a rotary-to-linear motion converter 322 is interleaved or provided between the motor shaft of motor 321 and the nut of helical spline mechanism 320, for converting rotary motion of the motor shaft in a normal-rotational direction or in a reverse-rotational direction into axial movement of the nut of helical spline mechanism 320. In the shown embodiment, motor 321 is installed on the cylinder head of engine 1. In lieu thereof, motor 321 may be installed on camshaft timing pulley 14. In the case of motor 321 installed on the cylinder head of engine 1, a bearing has to be attached to the rotary-to-linear motion converter 322. In this case, rotary motion of the motor shaft of motor 321 is converted into axial movement (linear motion) of the nut of helical spline mechanism 320 through the bearing. With the previously-noted arrangement of FIG. 3, the variable valve timing control function of the VTC system can be achieved or realized simultaneously with the start of cranking, by electrically controlling motor 321. In more detail, the actual relative phase of camshaft 310 to camshaft timing pulley 14 can be controlled or adjusted to a desired value by controlling rotary motion (or an applied current value) of motor 321 by way of closed-loop control responsively to the sensor signal from camshaft sensor 16. As a reversible motor that creates axial movement of the nut of helical spline mechanism 320, the VTC system may use a D. C. motor, a stepping motor, a synchronous motor with a permanent magnet, or the like. In the case of the use of rotary-to-linear motion converter 322 having a design speed reduction ratio set to a comparatively great value, it is more desirable or preferable that the VTC mechanism is conditioned in its maximum phase-retard state, in advance, before the start of cranking, so that the VTC mechanism can be kept at the maximum phase-retard state, even under a condition where a drop in battery voltage occurs during the cranking period.

In the engine control system of the compression ignition engine of the embodiment, it is necessary to control relative phase of camshaft 310 to camshaft timing pulley 14 by means of the VTC mechanism during the cranking period. Therefore, even at very low engine speeds, substantially corresponding to zero, the engine control system uses information concerning the actual relative phase of the VTC mechanism. For the reasons discussed above, the engine control system of the compression ignition engine of the embodiment uses the high-precision camshaft sensor 16 having a high detection accuracy at which camshaft sensor 16 is able to detect the angular phase of camshaft 310 (in other words, the operating state of intake valve 9) even at very low engine speeds, substantially corresponding to zero.

Referring to FIG. 4, there is shown the detailed construction of high-precision camshaft sensor 16. As shown in FIG. 4, camshaft sensor 16 is comprised of a toothed portion 330 attached to camshaft 310, a bridge circuit having magnetic resistance elements 331 located close to the toothed portion 330, and a magnet 332. The bridge circuit, having magnetic resistance elements 331, is disposed between the toothed portion 330 and magnet 332. The strength of magnetic flux 333 produced by magnet 332 varies depending on the relative position of each tooth of toothed portion 330 to magnet 332. When an electric resistance of each of magnetic resistance elements 331 varies owing to a change in magnetic flux 333, a change in electric voltage in the bridge circuit occurs. One pair of diagonally opposite corners of the bridge circuit is connected to a first one of two input terminals of each of a difference circuit (DIFF circuit) 334 and a summation circuit (SUM circuit) 335, whereas the other pair of the bridge circuit is connected to the second input terminal of each of DIFF circuit 334 and SUM circuit 335. The output terminal of DIFF circuit 334 generates a differential signal (simply, a DIFF signal), whereas the output terminal of SUM circuit 335 generates a summation signal (simply, a SUM signal). Based on the DIFF signal and the SUM signal, it is possible to detect or determine whether the toothed portion 330 of camshaft 310 is rotating or stationary.

Referring to FIG. 5, there is shown the detailed construction of another type of high-precision camshaft sensor 16. As shown in FIG. 5, camshaft 310 has a cam 343. A rotor 341 is installed on or fixedly connected to the left-hand side (the crankshaft side or the camshaft timing pulley side) of a VTC mechanism 342, whereas a rotor 344 is installed on or fixedly connected to the right-hand side of cam 343 of camshaft 310. Rotor 341 has a leaf 340, which is integrally connected onto or integrally formed with the outer periphery of rotor 341 and whose radial height varies according to its rotational directions. In a similar manner, rotor 344 has a leaf 345, which is integrally connected onto or integrally formed with the outer periphery of rotor 344 and whose radial height varies according to its rotational directions. The radial height of leaf 340 is detected by a gap sensor 346, and then the detected radial height of leaf 340 is converted into an analogue voltage signal (output signal from gap sensor 346). In a similar manner, the radial height of leaf 345 is detected by a gap sensor 350, and then the detected radial height of leaf 345 is converted into an analogue voltage signal (output signal from gap sensor 350). The analogue voltage signal from gap sensor 346 is converted into an angular signal via a signal converter (or an arithmetic circuit) 347, whereas the analogue voltage signal from gap sensor 350 is converted into an angular signal via a signal converter (or an arithmetic circuit) 349. These angular signals are input from signal converters 347 and 349 into a cam rotation angle arithmetic unit 348. Within cam rotation angle arithmetic unit 348, the valve timing of intake valve 9, for example, intake valve closure timing IVC is calculated based on the angular signal from signal converter 347 and the angular signal from signal converter 349, and thereafter the calculated signal (IVC signal) is generated from the output terminal of cam rotation angle arithmetic unit 348.

Referring to FIG. 6, there is shown the relationship between a crank angle and a camshaft sensor signal output value of the camshaft sensor of FIG. 5, exactly, the crank-angle versus first gap sensor voltage signal characteristic of gap sensor 346 and the crank-angle versus second gap sensor voltage signal characteristic of gap sensor 350. On the basis of the output voltage signal from the 1^(st) gap sensor 346, a reference cam rotation angle signal (see the output voltage signal characteristic indicated by the solid line in FIG. 6) is generated within signal converter 347. On the basis of the output voltage signal from the 2^(nd) gap sensor 350, a camshaft rotation angle signal, that is, a cam rotation angle signal (see the output voltage signal characteristic indicated by the broken line in FIG. 6) is generated within signal converter 349. Cam rotation angle arithmetic unit 348 arithmetically calculates or computes a cam timing controlled variable (i.e., a cam timing advance amount) based on the generated reference cam rotation angle signal and the generated camshaft rotation angle signal. Thus, intake valve closure timing IVC of intake valve 9 can be controlled in response to the output signal, generated from cam rotation angle arithmetic unit 348 and representative of the calculated cam timing advance amount.

In the case that the engine control system of the compression ignition engine of the embodiment uses camshaft sensor 16, which is comprised of the toothed portion 330, the bridge circuit having magnetic resistance elements 331, and magnet 332 and shown in FIG. 4, the sensor signal from crank angle sensor 5 of FIG. 1 is utilized as the reference cam rotation angle signal of the crank-angle versus camshaft sensor signal characteristic as shown in FIG. 6. As the camshaft rotation angle signal (the cam rotation angle signal), the DIFF signal and the SUM signal generated from camshaft sensor 16 of FIG. 4 are utilized. After the toothed portion 330 of camshaft 310 starts to rotate, the DIFF signal and the SUM signal are generated from camshaft sensor 16. Concretely, the DIFF signal and the SUM signal are outputted each time each of teeth of the toothed portion 330 approaches close to magnet 332. A specified one of teeth of the toothed portion 330, corresponding to the zero camshaft rotation angle of camshaft 310, is cut out, to provide a reference point of camshaft rotation angle. Thus, the number of pulses of each of the DIFF signal and the SUM signal can be counted from the cut-out portion serving as the reference point (the reference camshaft angular position). Based on the counted value of pulses of each of the DIFF signal and the SUM signal, the camshaft rotation angle signal is generated.

Referring now to FIG. 7, there are shown the timings IVO and IVC of intake valve 9 and the timings EVO and EVC of exhaust valve 10 during normal engine operation in a four-stroke-cycle compression ignition engine, such as a four-stroke-cycle Diesel engine. Exhaust valve 10 starts to open substantially at minus 180 degrees of crank angle at the beginning of exhaust stroke. The timing where exhaust valve 10 starts to open is called as “exhaust valve open timing EVO”. Exhaust valve 10 starts to close at the end of exhaust stroke. The timing where exhaust valve 10 starts to close is called as “exhaust valve closure timing EVC”. On the other hand, intake valve 9 starts to open at a timing value near 0° crank angle at the beginning of intake stroke. The timing where intake valve 9 starts to open is called as “intake valve open timing IVO”. Intake valve 9 starts to close at a timing value near BDC (corresponding to 180 degrees of crank angle) at the end of intake stroke. The timing where intake valve 9 starts to close is called as “intake valve closure timing IVC”. Diesel fuel (fuel oil) is sprayed or injected into the cylinder at the end of compression stroke. Self-ignition of the sprayed fuel occurs before or after TDC (corresponding to 360 degrees of crank angle). When intake valve closure timing IVC is phase-advanced from BDC, the quantity of gas (air) charged in the cylinder tends to reduce, thus resulting in a reduced effective compression ratio. Conversely when intake valve closure timing IVC is phase-retarded from BDC, reflux of gas (air), charged into the cylinder once, back to induction system 20 occurs, and thus the mass of gas charged into the cylinder is reduced, thereby resulting in a reduced effective compression ratio.

Referring now to FIG. 8, there are shown the timings IVO and IVC of intake valve 9 and the timings EVO and EVC of exhaust valve 10 during normal engine operation in a two-stroke-cycle compression ignition engine, such as a two-stroke-cycle Diesel engine. One operating cycle of events, namely the intake and compression strokes as well as the power and exhaust strokes, is completed for every crankshaft revolution (360 degrees of-crank angle). During the first 180 degrees crank angle range (in a crank angle range from 0° to 180°), the intake and compression strokes are produced. During the subsequent 180 degrees crank angle range (in a crank angle range from 180°to 360°), the power and exhaust strokes are produced. Self-ignition of the sprayed fuel occurs before TDC, corresponding to 180 degrees of crank angle. The valve-opening action of intake valve 9 and the valve-opening action of exhaust valve 10 are performed once for each crankshaft revolution. Thus, in the two-stroke-cycle engine, in FIG. 1 camshaft timing pulley 14 is driven by the crankshaft at the same revolution speed of crankshaft 2. The other structure of the two-stroke-cycle compression ignition engine is similar to that of the four-stroke-cycle compression ignition engine. In the two-stroke-cycle compression ignition engine, when intake valve closure timing IVC is brought closer to BDC, gas (air) charged in the cylinder is compressed under a condition where the mass of the charged gas is great, thus increasing or rising the effective compression ratio. On the contrary, when intake valve closure timing IVC is phase-retarded with respect to BDC, for the same internal pressure in induction system 20, there is a tendency for the quantity of gas (air) charged into the cylinder to be reduced, thus lowering or decreasing the effective compression ratio.

Referring to FIG. 9, there is shown a phase-control characteristic obtained by means of the VTC system of the compression ignition engine of the embodiment during a cranking and starting period. When the latest up-to-date informational data of engine speed Ne, determined based on the sensor signal from crank angle sensor 5, indicates a very low engine speed substantially corresponding to zero during the starting period of engine 1, or when the latest up-to-date informational data of engine temperature Te, determined based on the sensor signal from engine temperature sensor 15, indicates a low engine temperature value less than or equal to 40° C. with the ignition switch turned ON, the processor of ECU 6 determines that engine 1 is in a cold starting state. At the time ta at which the engine starter becomes energized (ON), the phase of variable valve actuation mechanism 13 (hereinafter referred to as “VTC phase”) is retarded from a phase suited to normal engine operation of engine 1. As can be seen from the phase-change characteristic curves of FIG. 10 attained by the intake-valve VTC system, intake valve closure timing IVC is phase-retarded from BDC during cranking. As a result, the effective compression ratio is lowered, thus reducing the work of compression. This contributes to the increased cranking speed and enhanced startability. Suppose that the VTC phase has been retarded in advance, before the start of cranking. In such a case, starter 7 becomes energized (ON) at once without VTC phase-retard control.

Returning to FIG. 9, at the time tb at which the cranking speed begins to exceed 400 rpm, the VTC mechanism is controlled so that the VTC phase (that is, intake valve closure timing IVC) is brought closer to the phase corresponding to the maximum phase-advance state. Therefore, as can be seen from the intake-valve phase changes shown in FIG. 10, intake valve closure timing IVC is brought closer to a timing value near BDC. As a result, the effective compression ratio is raised, and thus in-cylinder gas temperature becomes high. Thereafter, at the time tc at which the VTC phase becomes the phase corresponding to the maximum phase-advance state, fuel injection starts. At the time tc, by virtue of the high effective compression ratio and high temperature gas, the injected fuel is certainly combusted. Owing to the combustion energy, a rapid rise in engine speed Ne occurs. At the same time, in order to adjust the fuel-injection amount, the electronic fuel injection system is controlled by means of ECU 6 of FIG. 1, such that engine speed Ne is maintained at a specified idling speed for example 600 rpm. In this manner, in the engine control system of the compression ignition engine of the embodiment, the VTC phase (i.e., intake valve closure timing IVC) can be controlled temporarily to the phase corresponding to the phase-advance state during the starting period. Therefore, suppose that the engine control system of the compression ignition engine of the embodiment, capable of varying the effective compression ratio by a change in the IVC phase (a change in intake valve closure timing IVC), is applied to a fixed compression-ratio compression-ignition internal combustion engine (of a low geometrical compression ratio) in which intake valve closure timing IVC is generally controlled to a timing value phase-retarded from BDC by way of feedback control after engine warm-up. It is possible to adjust the effective compression ratio temporarily to a high ratio by controlling intake valve closure timing IVC to a timing value phase-advanced from BDC during the starting and warm-up period, as if the geometrical compression ratio has been varied to a high ratio for example by means of a multi-link variable compression ratio device. That is to say, the variable valve operating system (variable valve actuation mechanism 13) incorporated in the compression ignition engine of the embodiment, capable of varying the effective compression ratio by phase-changing intake valve closure timing IVC, has a variable compression ratio function just like a multi-link variable compression ratio device (or a multi-link piston crank mechanism) capable of varying a geometrical compression ratio, defined-as a ratio (V1+V2)/V1 of the full volume (V1+V2) existing within the engine cylinder and combustion chamber with the piston at BDC to the clearance-space volume (V1) with the piston at TDC, by varying a piston stroke characteristic (at least one of the piston TDC position and the piston BDC position). Generally, the low-geometrical-compression-ratio compression-ignition engine (the fixed geometrical-compression-ratio engine) itself has merits, namely light weight and reduced cost. In the case of a combination of the low-geometrical-compression-ratio compression-ignition engine and the variable valve operating system (variable valve actuation mechanism 13) incorporated in the compression ignition engine of the embodiment capable of varying the effective compression ratio by phase-changing intake valve closure timing IVC, there are several advantages, that is, light-weight, simple construction, and reduced cost. As discussed above, the geometrical compression ratio often denoted by Greek letter “ε” is generally defined as a ratio (V1+V2)/V1 of the full volume (V1+V2) existing within the engine cylinder and combustion chamber with the piston at BDC to the clearance-space volume (V1) with the piston at TDC. On the other hand, the effective compression ratio denoted by Greek letter “ε′” is generally defined as a ratio of the effective cylinder volume corresponding to the maximum working medium volume to the effective clearance volume corresponding to the minimum working medium volume. These two compression ratios ε and ε′ are thermodynamically distinguished from each other.

Hereupon, it is necessary to care the fact that the quantity of air charged into the cylinder of engine 1 changes depending on intake valve closure timing IVC. When intake valve closure timing IVC is retarded, the quantity of air charged into engine 1 becomes small. Therefore, it is desirable to properly control the fuel-injection amount, fully taking into account the intake valve closure timing IVC. For this reason, the fuel-injection amount is compensated for responsively to at least sensor signals from camshaft sensor 16 and air flow sensor 17 in addition to engine speed Ne and engine load (e.g., the amount APS of depression of the accelerator pedal), thereby preventing or suppressing the generation of soot.

Thereafter, as can be seen from the A characteristic curve (indicated by the solid line in FIG. 9), at the time td at which a stable combustion state has been reached, for example, when the engine temperature (engine coolant temperature) exceeds 60° C., the VTC phase is controlled or adjusted from the phase corresponding to the maximum phase-advance state in the phase-retard direction (that is, toward a phase suited to the normal engine operation of engine 1). In lieu of the VTC phase control based on the A characteristic curve of FIG. 9, the VTC phase control based on the B characteristic curve indicated by the broken line in FIG. 9 may be applied. According to the VTC phase control based on the B characteristic curve of FIG. 9, during the warm-up time period tc-td from tc to td, intake valve closure timing IVC may be retarded gradually to the phase suited to the normal engine operation of engine 1, depending on the combustion stability. As can be seen from the phase-change characteristic curves of FIG. 10 attained by the intake-valve VTC system, the VTC phase suited after completion of engine warm-up, is set to an intermediate phase, which is advanced from the VTC phase suited for the early stage (see the time period ta-tb in FIG. 9) of engine cranking and retarded from the maximum phase-advanced VTC phase of the starting and warm-up period (in particular, see the warm-up time period tc-td in FIG. 9). By way of the proper setting of the VTC phase (i.e., the previously-noted intermediate phase) suited after completion of engine warm-up, it is possible to properly lower the effective compression ratio to such an extent that does not cause any trouble in combustion, thereby effectively lowering the work of compression and ensuring a reduction in mechanical friction loss of engine 1, a decrease in fuel consumption rate (improved fuel economy), and reduced NOx (nitrogen oxides) emissions. Additionally, by virtue of the properly lowered effective compression ratio after engine warm-up, it is possible to effectively reduce a peak value of combustion pressure, thereby ensure a reduction in combustion noise and vibrations. Suppose that the engine control system of the compression ignition engine of the embodiment is applied to a hybrid vehicle employing an automatic engine stop-restart system frequently executing engine stop and restart operation even after completion of engine warm-up. It is possible to reconcile the enhanced startability and the reduced electric power consumption, by virtue of the properly lowered effective compression ratio after engine warm-up. From the time td after completion of engine warm-up, intake valve closure timing IVC is optimally controlled usually by way of closed-loop control depending on engine operating conditions such as engine speed Ne and engine load (accelerator-pedal depression amount APS).

Referring now to FIG. 11, there is shown the starting-period VTC control routine executed within ECU 6 incorporated in the engine control system of the compression ignition engine of the embodiment. The routine of FIG. 11 is executed as time-triggered interrupt routines to be triggered every predetermined time intervals for example 10 milliseconds.

After an ignition switch (an engine key switch) is turned ON at step 360, a check or a determination for the current VTC phase is made through step 361. Actually, at step 362 just after step 361, the processor of ECU 6 executes a comparative check for the current VTC phase (that is, the current intake valve closure timing IVC) with respect to a first predetermined phase angle. More concretely, when step 362 determines that the current VTC (i.e., the latest up-to-date informational data of intake valve closure timing IVC) is phase-retarded with respect to the first predetermined phase angle, the routine proceeds to step 363. Conversely when step 362 determines that the current intake valve closure timing IVC is not phase-retarded with respect to the first predetermined phase angle, the routine proceeds to step 364.

At step 363, starter 7 becomes energized (ON).

At step 364, starter 7 becomes energized (ON).

Subsequently to step 364, step 365 occurs to initiate VTC phase-retard control for variable valve actuation mechanism 13 (the VTC mechanism). In the control routine of FIG. 11, step 365 (that is, VTC phase-retard control) is executed just after step 364 (that is, starter energizing operation). In lieu thereof, step 365 (VTC phase-retard control) may be executed just before step 364 (starter energizing operation).

At step 367, a comparative check similar to step 362 is made again to determine whether the current VTC phase (that is, the current intake valve closure timing IVC) is phase-retarded with respect to the first predetermined phase angle. When the answer to step 367 is affirmative (YES), that is, when the current intake valve closure timing IVC has been phase-retarded with respect to the first predetermined phase angle, the routine proceeds to step 368. Conversely when the answer to step 367 is negative (NO), the routine returns to step 365. The return from step 367 to step 365 is repeatedly executed until the actual intake valve closure timing IVC has been retarded with respect to the first predetermined phase angle. In other words, after a specified time period has expired from the initial execution of step 367, the routine shifts from step 367 to step 368. The current timing value of intake valve closure timing IVC, needed for the comparative check of step 367, is detected or determined based on the sensor signal from camshaft sensor 16.

At step 368, a check is made to determine whether the latest up-to-date informational data of engine speed Ne, determined based on the sensor signal from crank angle sensor 5, is greater than or equal to a first predetermined speed value such as 400 revolutions per minute. When the answer to step 368 is affirmative (YES), that is, when the current engine speed is greater than or equal to the first predetermined speed value (e.g., 400 rpm), the routine proceeds from step 368 to step 369. Conversely when the answer to step 368 is negative (NO), step 368 is repeatedly executed, until the current engine speed exceeds the first predetermined speed value owing to a rise in cranking speed.

At step 369, VTC phase-advance control for variable valve actuation mechanism 13 (the VTC mechanism) is executed. After step 369, step 370 occurs.

At step 370, a comparative check is made to determine whether the current VTC phase (that is, the current intake valve closure timing IVC) is phase-advanced with respect to a second predetermined phase angle. When the answer to step 370 is affirmative (YES), that is, when the current intake valve closure timing IVC has been phase-advanced with respect to the second predetermined phase angle, the routine proceeds to step 371. Conversely when the answer to step 370 is negative (NO), step 370 is repeatedly executed, until the current intake valve closure timing IVC has been phase-advanced with respect to the second predetermined phase angle by way of the VTC phase-advance control.

At step 371, fuel injection starts. At this time, the fuel-injection amount is compensated for responsively to at least the sensor signal from camshaft sensor 16 and the sensor signal (representative of the quantity of air charged into the cylinder) from air flow sensor 17 in addition to engine speed Ne and engine load (e.g., accelerator-pedal depression amount APS). After step 371, step 372 occurs.

At step 372, a check is made to determine whether the latest up-to-date informational data of engine speed Ne (i.e., the current engine speed) is greater than or equal to a second predetermined speed value such as 500 revolutions per minute. When the answer to step 372 is affirmative (YES), that is, when the current engine speed is greater than or equal to the second predetermined speed value (e.g., 500 rpm), the routine proceeds from step 372 to step 373. Conversely when the answer to step 372 is negative (NO), step 372 is repeatedly executed, until the current engine speed exceeds the second predetermined speed value.

At step 373 a check is made to determine whether the latest up-to-date informational data of engine temperature, determined based on the sensor signal from engine temperature sensor 15, is greater than or equal to a predetermined temperature value (a temperature threshold) such as 60° C. When the answer to step 373 is affirmative (YES), that is, when the current engine temperature is greater than or equal to the predetermined temperature value (,e.g., 60° C.), the routine proceeds from step 373 to step 374. Conversely when the answer to step 373 is negative (NO), step 373 is repeatedly executed, until the current engine temperature exceeds the predetermined temperature value owing to an engine temperature rise after the fuel injection operation.

At step 374, VTC phase-retard control for variable valve actuation mechanism 13 (the VTC mechanism) is executed so that the actual VTC phase (that is, the actual intake valve closure timing IVC) is retarded towards a phase suited to the normal engine operation of engine 1. After step 374, step 375 occurs.

At step 375, a check is made to determine whether the current VTC phase (i.e., the actual intake valve closure timing IVC) is brought closer to the desired phase (the desired timing) suited for the normal engine operation and determined based on the up-to-date informational data of engine speed Ne and engine load APS, by way of closed-loop control for the VTC phase. In this manner, through step 375, the intermediate VTC phase control suited for the normal engine operation is executed within an intermediate phase-angle range phase-advanced from the maximum phase-retarded VTC phase and retarded from the maximum phase-advanced VTC phase.

According to the control routine of FIG. 11, only during the early stage (see the time period ta-tb in FIG. 9) of engine cranking, intake valve closure timing IVC is retarded from BDC, and thereafter intake valve closure timing IVC is adjusted (phase-advanced) to a timing value near BDC so that engine 1 is operated in a high effective compression ratio mode. After this, immediately when engine speed Ne reaches and exceeds the second predetermined speed value, for example, 500 rpm, the processor of ECU 6 determines that the engine starting operation has been completed. After completion of engine starting operation, in other words, after completion of engine warm-up, intake valve closure timing IVC is,retarded by a predetermined phase angle Δ (corresponding to a phase difference between the actual intake valve closure timing IVC and its desired value) from BDC. In order to properly lower the effective compression ratio after completion of engine starting operation, in the case of the four-stroke-cycle compression ignition engine, intake valve closure timing IVC may be phase-advanced from BDC.

In the engine stopped state, as can be supposed, there are two cases, namely one being a case that intake valve closure timing IVC (VTC phase) has already been set to the maximum phase-retard timing under the engine stopped state, and the other being a case that intake valve closure timing IVC (VTC phase) is controlled to the maximum phase-retard timing simultaneously with the turned-ON operation of the ignition switch. Therefore, in the case that intake valve closure timing IVC (VTC phase) has to be controlled to the maximum phase-retard timing simultaneously with the ignition switch turned-ON operation, through steps 365 and 367 the current intake valve closure timing IVC (or the actual IVC) is detected or determined based on the sensor signal from camshaft sensor 16, and then a phase-angle difference (a deviation or an error signal) between the detected actual intake valve closure timing IVC and the desired value (e.g., the maximum phase-retard timing) is determined. In order to adjust the phase-angle difference to zero, VTC phase control (IVC control) is performed. On the contrary, in the case that intake valve closure timing IVC (VTC phase) has already been set to the maximum phase-retard timing under the engine stopped state, the routine of FIG. 11 flows from step 362 (only the comparative check for the current VTC phase (the current IVC) with respect to the first predetermined phase angle) through step 363 (starter energizing operation) quickly to step 368, bypassing steps 365 and 367.

In the starting-period VTC control system of FIG. 11, after step 371 (the fuel-injection starting step), (i) the first check for an engine speed rise above the second predetermined speed value (e.g., 500 rpm) and (ii) the second check for an engine temperature rise above the predetermined temperature value (e.g., 60° C.) are both made. In lieu thereof, only the first check may be executed, without executing the second check. In such a case, immediately when an engine speed rise above the second predetermined speed value is detected, the routine advances from step 372 quickly to step 374 without executing a check for an engine temperature rise, so that the actual intake valve closure timing IVC can be quickly retarded by the predetermined phase angle Δ, corresponding to the phase difference between the actual intake valve closure timing IVC and the desired value, corresponding to the intermediate VTC phase angle suited for the normal engine operation. This is because, in the case of engine speeds above the second predetermined speed value (e.g., 500 rpm) during the last stage of engine starting, the quantity of heat generated by combustion of the injected fuel tends to be greater than the quantity of heat radiated from the cylinder wall, even under a comparatively low engine-temperature condition, and thus stable combustion can be ensured. As soon as the processor of ECU 6 determines, based on only the check for the engine speed rise above the second predetermined speed value (500 rpm), that a stable combustion state has been reached and achieved, the routine proceeds to step 375 to initiate the intermediate VTC phase control suited for the normal engine operation.

In the case that intake valve closure timing IVC is not set to the maximum phase-retard timing in the engine stopped state, through step 365 of the control routine of FIG. 11, the VTC phase-retard control, in other words, IVC phase-control to the maximum phase-retard timing is executed. As discussed previously, the quantity of air charged into the cylinder changes depending on intake valve closure timing IVC. Thus, fully taking into account the air-fuel mixture (A/F) ratio, the fuel-injection amount must be varied responsively to the VTC phase change (i.e., IVC phase change). Actually, the mass of fuel, injected from fuel injection valve 4, is controlled or changed depending on various factors, that is, the quantity Qa of air entering the engine cylinder, measured by air flow sensor 17, the accelerator-pedal depression amount APS, and engine speed Ne detected by crank angle sensor 5. In addition to the aforementioned sensor signals, preferably, the state of EGR valve 19 (i.e., the EGR valve opening) and the turbo-charging state (e.g., boost pressure) of turbo charger 18 may be taken into account for determining the mass of the injected fuel and injection timing. As a matter of course, in the case of the occurrence of a change in VTC phase, that is, a change in intake valve open timing IVO as well as a change in intake valve closure timing IVC, it is necessary to properly change the fuel injection timing as well as fuel-injection amount. For the reasons discussed above, the sensor signal input from camshaft sensor 16 into the input interface of ECU 6 is important to execute the VTC phase control (that is, IVC phase control) and also to execute the electronic fuel injection control. For instance, in the case of one-stroke injection of the Diesel engine, a fuel injection pattern is classified into a pilot-injection area, a main-injection area, an after-injection area, and a post-injection area. The fuel injection pattern changes depending on engine operating conditions. In step 371 of the flow chart shown in FIG. 11, the fuel injection pattern is given as a function of intake valve closure timing IVC. By the use of the predetermined or preprogrammed function representative of the relationship between the fuel injection pattern and intake valve closure timing IVC, a change in intake valve closure timing IVC can be remarkably reflected as a change in the fuel injection pattern (containing a fuel injection amount and the number of fuel injections).

Instead of executing (i) a check for the necessity for VTC phase retard and (ii) VTC phase-retard control (see the flow from step 361 via steps 362 and 364 to step 365 in FIG. 11) during the early stage of the engine starting period, the control routine as shown in FIG. 11 may be somewhat modified, so that the VTC phase control (i.e., the IVC control) to a desired phase (i.e., a desired standby timing suited for the early stage of engine cranking) is executed for each transition to an engine stopped state. In this case, it is possible to easily adjust intake valve closure timing IVC to a desired standby timing spaced apart from BDC for each transition to an engine stopped state, by slightly modifying step 365 of FIG. 11. Concretely, at the modified step 365, a control command signal (or a drive signal or an electric signal) is applied to the actuator (or the electrically-controlled actuator means) of the VTC mechanism during the early stage of the engine stopping period in such a manner as to control or adjust the intake valve closure timing IVC to the desired standby timing spaced apart from BDC, that is, to the desired phase-change state remarkably phase-retarded or phase-advanced from BDC. After completion of the IVC phase-adjustment to the desired standby timing, at the last stage of the engine stopping period, an engine stop signal is output. By the use of the modified step 365, it is possible to efficiently achieve the IVC phase-adjustment to the desired standby timing spaced apart from BDC during the engine stopping period rather than the engine starting period. This eliminates (i) a check for the current VTC phase and (ii) VTC phase-retard control during the early stage of the engine starting period. This contributes to the shortened engine starting time.

The combustion stability of engine 1 is affected by various control parameters, namely, engine temperature, fuel property (for example, cetane value), intake air temperature (charge air temperature), residual gas ratio, EGR rate, boost pressure, and the like. In the engine control system of the compression ignition engine of the embodiment, it is preferable that a plurality of control parameters, directly participating in combustion stability, are detected and intake valve closure timing IVC is controlled, fully taking into account the detected control parameters. As the control parameters, directly participating in combustion stability, the following parameters are exemplified.

-   (1) In-cylinder pressure; -   (2) Vibrations of cylinder head, caused by gas vibrations of     controlled or uncontrolled burning -   (3) Rotational-speed fluctuations in crankshaft -   (4) Ionic current arising from combustion -   (5) Emission intensity of flame     A threshold value for each of the above-mentioned control     parameters, suited for engine speed and engine load after engine     warm-up, is experimentally measured and determined beforehand. Thus,     it is possible to determine, based on the comparison result of the     detected value of the control parameter and its threshold value,     whether or not a stable combustion state has been reached. Based on     such a decision result concerning a stable/unstable combustion     state, it is possible to prevent forcible phase-advance of intake     valve closure timing IVC during the starting period, thus enabling     intake valve closure timing IVC to timely control in the     phase-retard direction. At this time, it is necessary to     experimentally measure and determine the threshold value of each of     the control parameters in the VTC phase-advance state with a target     engine, beforehand. The threshold values for these control     parameters can be used for fuel-injection amount control and     fuel-injection timing control in real time.

Referring now to FIG. 12, there is shown the disassembled view of the electronically-controlled hydraulically-operated rotary vane type VTC mechanism, capable of executing the VTC phase-retard control of steps 365 and 374 and VTC phase-advance control of step 369. As can be seen from the disassembled view of FIG. 12, a hydraulically-operated vane body 22 of a four-blade vane unit 105 is fixedly connected or bolted to the shaft end of an intake camshaft 200 having intake cam 11, by means of a center bolt (or a vane mounting bolt) 21. Camshaft timing pulley 14 is formed with a sprocket 103, which serves as a rotary member driven by the engine crankshaft via a timing chain 131 (see FIG. 14A). Camshaft timing pulley 14 is fixedly connected to a substantially cylindrical hydraulic housing 23 integrally formed with four partition wall portions (simply, four shoes), each protruding radially inwards from and integrally formed with the inner periphery of the cylindrical housing. Vane body 22 is operably accommodated in hydraulic housing 23. The front end of vane body 22 is hermetically covered in a fluid-tight fashion by a front cover 24. Vane body 22 is formed integral with four vanes. Application of hydraulic pressure to one sidewall of each of the vanes causes relative rotary motion of vane body 22 to housing 23, thereby resulting in a phase difference (a change in relative phase) between vane body 22 and housing 23. Thus, it is possible to vary intake valve closure timing IVC during operation of engine 1 by controlling the phase difference between vane body 22 and housing 23. In the case of the electronically-controlled hydraulically-operated rotary vane type VTC mechanism shown in FIG. 12, intake valve open timing IVO varies simultaneously with a change in intake valve closure timing IVC.

As clearly shown in FIG. 12, two rows of return springs (biasing means) 25, 25 are disposed between one sidewall surface of each of the vanes and a stopper surface of each of the partition wall portions of housing 23. In total, eight return springs 25 are arranged in housing 23. Spring forces of return springs 25 permanently bias vane body 22 to cause relative rotary motion of vane body 22 to housing 23 in a clockwise direction, that is, in a phase-advance direction of camshaft 200. Alternatively, return springs 25 may be arranged in housing 23, so that spring forces of the return springs permanently bias vane body 22 to cause relative rotary motion of vane body 22 to housing 23 in a counterclockwise direction, that is, in a phase-retard direction of camshaft 200. Front cover 24 is secured to or fixedly connected to housing 23 by means of mounting bolts 117. Although it is not clearly shown in FIG. 12, front cover 24 is formed with an air bleeder hole 128. As described later in reference to FIGS. 14A-14C, to create rotary motion of each of the vanes of vane body 22 relative to housing 23, for phase advance, working fluid (hydraulic oil) is supplied into a variable-volume phase-advance hydraulic chamber 30 through a phase-advance hydraulic line 32 and a phase-advance oil hole 106. Conversely for the purpose of phase retard, working fluid (hydraulic oil) is supplied into a variable-volume phase-retard hydraulic chamber 31 through a phase-retard hydraulic line 33 and a phase-retard oil hole 107. The four vane blades of vane body 22 and housing 23 cooperate with each other to define four variable-volume phase-retard chambers 31 and four variable-volume phase-advance chambers 30. In the shown embodiment, the two rows of springs (25, 25) are disposed in the associated phase-advance chamber 30.

Phase-advance hydraulic line 32 and phase-retard hydraulic line 33 are formed or defined in intake camshaft 200 shown in FIG. 12. Working fluid discharged from the engine oil pump, provided for lubricating oil supply into the engine, or working fluid discharged from separate electric-motor driven hydraulic oil pump 302, is delivered through a phase-advance oil-delivery groove 35 and a phase-retard oil-delivery groove 34 into the respective hydraulic lines 32 and 33. Phase-advance oil-delivery groove 35 and phase-retard oil-delivery groove 34 are located in a cam journal bearing portion 108. The shaft end of intake camshaft 200 has a female screw-threaded portion 118 into which center bolt 21 is screwed. Working fluid flow of the previously-noted hydraulic circuit for the hydraulically-operated rotary vane type VTC mechanism is controlled by an oil control valve 39 whose operation is hereunder described in reference to FIGS. 13A-13C. Oil control valve 39 is comprised of an electromagnetic solenoid 40, a spool 41, and a spool-biasing spring 42. In FIGS. 13A-13C, a port denoted by “A” is connected to phase-advance hydraulic line 32, whereas a port denoted by “B” is connected to phase-retard hydraulic line 33. The axial position of solenoid 40 is controlled in response to a control signal (or a drive signal) applied from the output interface (or the drive circuit) of ECU 6 to the solenoid. That is, solenoid 40 serves as an actuator (a driving power source or an electrically-controlled actuator means) for the hydraulically-operated rotary vane type VTC mechanism.

As shown in FIG. 13A, when solenoid 40 is de-energized (OFF) and thus spool 41 is held at its spring-loaded position by the spring force of spool-biasing spring 42, hydraulic pressure in phase-advance hydraulic line 32 becomes high, while hydraulic pressure in phase-retard hydraulic line 33 becomes low. As a result of this, vane body 22 moves in the phase-advance direction, and then vane body 22 is held at an angular position corresponding to the maximum phase-advanced VTC phase, in other words, the maximum phase-advance intake valve closure timing substantially corresponding to BDC (that is, IVC≈BDC). The phase-advance state (IVC≈BDC) of the VTC mechanism, created by the spring-loaded axial position of spool 41, corresponds to an operating mode important to the starting and warm-up period of engine 1. Even in the presence of a failure in the engine control system such as a control signal line failure, it is possible to certainly start the engine 1 by the phase-advance state (IVC≈BDC) of the VTC mechanism, created by the spring-loaded axial position (the de-energized state) of spool 41.

As shown in FIG. 13B, when spool 41 moves axially leftwards against the spring force of spool-biasing spring 42 with solenoid 40 energized (ON) and then spool 41 is held at the leftmost spool position (viewing FIG. 13B), hydraulic pressure in phase-advance hydraulic line 32 becomes low, while hydraulic pressure in phase-retard hydraulic line 33 becomes high. As a result of this, vane body 22 moves in the phase-retard direction, and then vane body 22 is held at an angular position corresponding to the maximum phase-retarded VTC phase, in other words, the maximum phase-retard intake valve closure timing retarded from and spaced apart from BDC.

As shown in FIG. 13C, when spool 41 is held at a specified intermediate spool position, phase-advance hydraulic line 32 and phase-retard hydraulic line 33 are blocked by the two lands of spool 41. Hydraulic pressure in variable-volume phase-advance hydraulic chamber 30 and hydraulic pressure in variable-volume phase-retard hydraulic chamber 31 are held constant. That is, the VTC mechanism is held at its pressure-hold mode. As a result, the relative position of vane body 22 to housing 23 can be held at a desired angular position, in other words, at a balanced position of torque acting on vane body 22 due to the spring force of each of return springs 25 and torque acting on vane body 22 due to the differential pressure between phase-advance and phase-retard hydraulic chambers 30 and 31 defined on both sides of each of the vanes. Therefore, by properly controlling the axial position of spool 41, it is possible to hold intake valve closure timing IVC at an arbitrary timing value between the maximum phase-retard timing and the maximum phase-advance timing. The spool position control is executed by ECU 6 by way of closed-loop control (feedback control) based on the sensor signal from camshaft sensor 16. As discussed previously, during engine cranking, in particular, during the early stage (see the time period ta-tb in FIG. 9) of cranking, intake valve closure timing IVC is controlled to the maximum phase-retard timing, corresponding to the rightmost spool position shown in FIG. 13A. This results in the reduced work of compression, the increased cranking speed, and the ease of engine starting. Owing to the reduced work of compression, it is possible to easily crank the engine, in spite of setting of the applied electric current to starter 7 to a comparatively low level. This eliminates the necessity for an engine starter of a high torque capacity.

As set forth above, by way of the axial position control for spool 41, in other words, by way of applied current control for solenoid 40, as can be seen from the phase-change characteristic curves of FIG. 10, intake valve closure timing IVC can be controlled to an arbitrary timing value, ranging from the maximum phase-advance timing substantially corresponding to BDC to the maximum phase-retard timing retarded by approximately 40 degrees of crank angle from BDC. Intake valve open timing IVO also varies simultaneously with a change in intake valve closure timing IVC (see the intake-valve phase-change characteristic curves of FIG. 10). Approaching intake valve closure timing IVC closer to BDC results in a rise in effective compression ratio, thereby enhancing the startability of engine 1.

Additionally, during the cranking period, intake valve closure timing IVC is considerably phase-retarded from BDC, and as a result the work of compression is effectively reduced and a cranking speed increase occurs, thus ensuring enhanced startability. After completion of engine warm-up, the effective compression ratio is lowered by slightly retarding intake valve closure timing IVC from BDC, thus effectively reducing a fuel consumption rate after engine starting. Additionally, owing to the lowered effective compression ratio, an excessive rise in combustion temperature will be effectively suppressed, thus reducing NOx (nitrogen oxides) emissions.

Returning to FIG. 12, one of the four vane blades of vane body 22 has an axial bore that slidably accommodating therein a hydraulic lock piston 110. Piston 110 is arranged to selectively engaged with or disengaged from a seat 111 (having a lock-piston hole) of camshaft timing pulley 14, depending on engine operating conditions. With piston 110 in fitted-engagement with seat 111, vane body 22 is coupled to cam timing pulley 14, so that vane body 22 rotates together with cam timing pulley 14 during operation. For instance, when hydraulic pressure to be supplied to vane body 22 is insufficient due to a failure in separate electric-motor driven hydraulic oil pump 302 during a starting period, lock piston 110 is brought into engagement with seat 111 of camshaft timing pulley 14, thus constraining rotary motion (free rotation) of vane body 22 relative to the cylindrical housing 23 and consequently preventing the camshaft from rotating relative to the crankshaft.

As best seen from FIG. 14A, the position of fitted-engagement of piston 110 with seat 111, is set or designed to provide the maximum phase-advanced VTC phase, in other words, the maximum phase-advance intake valve closure timing substantially corresponding to BDC (i.e., IVC≈BDC). When engine 1 starts to rotate and thus hydraulic pressure acting on vane body 22 becomes high, piston 110 moves against the spring force of a piston return spring 112 in a direction disengaging of piston 110 from seat 111, under pressure of working fluid fed via phase-advance oil hole 106 and phase-retard oil hole 107. As a result, vane body 22 is uncoupled from camshaft timing pulley 14, thereby enabling vane body 22 to be controlled hydraulically.

As shown in FIG. 12, in addition to return springs 25, it is more preferable to further provide a torsion coil spring 120 between vane body 22 and front cover 24. There is no risk of interference between each of return springs 25 and torsion coil spring 120, since the installation position of each return spring 25 arranged in housing 23 differs from the installation position off torsion coil spring 120. A combination of return springs 25 and torsion coil spring 120 realizes a great magnitude of spring bias permanently biasing vane body 22 in a clockwise direction. Concretely, as appreciated from the disassembled view of FIG. 12, the left-hand hook end of torsion coil spring 120 is fitted into a torsion-spring hook insertion hole 122 bored in front cover 24, whereas the right-hand hook end of torsion coil spring 120 is fitted into a torsion-spring hook insertion hole 121 bored in vane body 22. In the same manner as return springs 25, a spring force of torsion coil spring 120 permanently biases vane body 22 to cause relative rotary motion of vane body 22 to housing 23 in a clockwise direction, that is, in a phase-advance direction of camshaft 200.

In FIG. 12, a member denoted by reference sign 104 is a positioning pin included in a positioning mechanism for the purpose of positioning between housing 23 and camshaft timing pulley 14 when assembling these component parts by means of bolts 117. The positioning means is effective to easily determine the specified angular position of housing 23 relative to camshaft timing pulley 14, in other words, the specified angular position of lock piston 110, which is slidably accommodated in the axial bore of vane body 22 circumferentially movable in housing 23 within limits, relative to the lock-piston hole of seat 111, when assembling the two component parts.

Referring to FIG. 14A, hydraulic housing 23 is driven by the engine crankshaft via a crankshaft timing pulley 132 and timing chain 131. In the case of the four-stroke-cycle engine, housing 23 is driven by the crankshaft at ½ the revolution speed of crankshaft 2. In the case of the two-stroke-cycle engine, housing 23 is driven by the crankshaft at the same revolution speed as crankshaft 2. As described previously, working fluid is supplied into variable-volume phase-advance hydraulic chamber 30 through phase-advance hydraulic line 32, whereas working fluid is supplied into variable-volume phase-retard hydraulic chamber 31 through phase-retard hydraulic line 33. When hydraulic pressure in phase-advance hydraulic chamber 30 is equal to or higher than that in phase-retard hydraulic chamber 31, phase-advance hydraulic chamber 30 is filled with working fluid (hydraulic oil). Under these conditions, vane body 22 is conditioned in its maximum phase-advance state (the maximum phase-advance angular position) shown in FIG. 14B. The valve-opening action and valve-closing action of intake valve 9 are made at the earliest timing with respect to a rotation angle of camshaft timing pulley 14, in other words, with respect to a crank angle. That is, intake valve closure timing IVC and intake valve open timing IVO are both set to their maximum phase-advance timings. When there is no application of hydraulic pressure to both of phase-advance hydraulic chamber 30 and phase-retard hydraulic chamber 31, the VTC phase (intake valve closure timing IVC and intake valve open timing IVO) is automatically controlled to the phase corresponding to the maximum phase-advance state shown in FIG. 14B by the spring forces produced by springs 25.

On the contrary, when hydraulic pressure in phase-retard hydraulic chamber 31 is higher than that in phase-advance hydraulic chamber 30, phase-retard hydraulic chamber 31 is filled with working fluid. Under these conditions, vane body 22 is conditioned in its maximum phase-retard state (the maximum phase-retard angular position) shown in FIG. 14C. The valve-opening action and valve-closing action of intake valve 9 are made at the latest timing with respect to a crank angle. That is, intake valve closure timing IVC and intake valve open timing IVO are both set to their maximum phase-retard timings.

As set forth above, by means of springs 25 disposed in respective phase-advance chambers 30, it is possible to automatically set intake valve closure timing IVC to the maximum phase-advance timing (i.e., IVC≈BDC) shown in FIG. 14B by the spring forces produced by springs 25, under a particular condition where there is no application of hydraulic pressure to both of phase-advance hydraulic chamber 30 and phase-retard hydraulic chamber 31. In the rotary vane type VTC mechanism of FIG. 12, return springs are exemplified in compression coil springs. Instead of using a compression coil spring as return spring 25, a tensile coil spring or a leaf spring may be used. By means of torsion coil spring 120 further provided in addition to springs 25, it is possible to automatically set intake valve closure timing IVC to the maximum phase-advance timing (i.e., IVC≈BDC) by the spring force produced by torsion coil spring 120, under a particular condition where there is no hydraulic-pressure application to both of phase-advance hydraulic chamber 30 and phase-retard hydraulic chamber 31.

In the case of the motor-driven spiral disk type VTC mechanism shown in FIG. 2, it is possible to automatically control or set the VTC phase, in particular, intake valve closure timing IVC to the maximum phase-advance state (i.e., IVC≈BDC) by a spring force produced by biasing means, even if there is no torque application to hysteresis member 316 owing to a failure in hysteresis motor 315. In this case, the biasing means is attached to helical spline mechanism 320 shown in FIG. 3, in such a manner as to enable automatic adjustment of the VTC phase, in particular, intake valve closure timing IVC to the phase corresponding to the maximum phase-advance state (i.e., IVC≈BDC) by the spring force produced by the biasing means attached to helical spline mechanism 320, even when reversible motor 321 is failed. By the provision of the biasing means, even in the presence of a failure in reversible motor 321, it is possible to certainly attain engine starting operation.

According to the VTC phase control shown in FIG. 9, by means of the variable valve actuation mechanism 13 (phase-change means), intake valve closure timing IVC is phase-retarded from BDC during cranking (see the time period ta-tb in FIG. 9), thereby reducing the work of compression. As shown in FIG. 15, another method of reducing the work of compression during cranking is to adjust or control intake valve closure timing IVC to a timing value phase-advanced from BDC. As can be seen from another phase-control characteristic of FIG. 15, as soon as engine cranking operation is initiated at the time point ta, the VTC phase is controlled to a phase advanced from BDC, and thus intake valve closure timing IVC of intake valve 9 is controlled to a timing value considerably phase-advanced from BDC. After the time tb at which the cranking speed begins to exceed 400 rpm, intake valve closure timing IVC is phase-retarded toward BDC. Thereafter, at the time tc, fuel injection starts. After this, at the time td when engine temperature Te exceeds a predetermined temperature value such as 60° C., intake valve closure timing IVC is retarded from BDC to a timing value suited to the normal engine operation of engine 1.

As another method of reducing the work of compression during cranking, a starting-period decompression device may be combined with phase change means or phase control means, such as the VTC mechanism, the VVL mechanism, the VEL mechanism or the like. The decompression device is provided to constantly open exhaust valve 10 during a cranking period, thereby permitting a reduction in the work of compression even when intake valve closure timing IVC of intake valve 9 has been phase-advanced to a timing value substantially corresponding to a phase-advance state. For instance, by pushing exhaust valve 10 downwards by means of an electromagnet, it is possible to slightly open exhaust valve 10, thus realizing a decompressing function. FIG. 16 shows the phase-control characteristic obtained by the combined system of the decompression device and the phase change means (the VTC mechanism).

Returning to FIG. 16, at the time ta, the starter becomes energized (ON) for engine cranking. At the same time, the VTC phase (intake valve closure timing IVC) is controlled to a phase substantially corresponding to a phase-advance state (=BDC), and additionally the decompression device is energized (ON) for maintaining exhaust valve 10 in its constantly-opened valve operating state (i.e., in a decompression mode). The decompression device may be energized (ON) before the time ta. As soon as the cranking speed begins to exceed 400 rpm at the time tb, the decompression device is de-energized (OFF) to inhibit the exhaust-valve decompression mode. And thus, the operating mode of exhaust valve 10 returns to a normal valve operating state. Thereafter, at the time tc, fuel injection starts. At the time tc1 after tc, an increase in fuel injection amount is inhibited or stopped, so as to control or maintain engine speed Ne to a specified idling speed for example 600 rpm. Thereafter, at the warm-up completion time point td at which combustion has been stabilized, for example, when engine temperature Te exceeds 60° C., the VTC phase is controlled from the phase substantially corresponding to the phase-advance state in the phase-retard direction (that is, toward a phase suited to the normal engine operation of engine 1). In the case of the VTC phase control shown in FIG. 16, the VTC mechanism is merely switched from one of the relatively phase-retarded position and the relatively phase-advanced position to the other. The VTC phase control system is simple in phase-control components. This contributes to the reduced cost of the VTC mechanism. Additionally, in the case of the combined system of the starting-period decompression device and the phase change means (the VTC mechanism), it is possible to completely control or adjust the work of compression to zero during the cranking and starting period. Therefore, the cranking process can be passed early, and the engine starting time can be shortened, thus reducing exhaust emissions such as soot.

On automotive vehicles, owing to a rapid engine torque rise, the vehicle body tends to vibrate undesirably. To avoid this, as can be seen from the phase-control characteristic of FIG. 17, the VTC phase is first controlled to a phase-retard state (considerably retarded from BDC) simultaneously with the start of cranking (see a rapid fall in the VTC phase from the time ta), and thereafter the VTC phase is gradually controlled moderately toward a phase corresponding to a phase-advance state (=BDC) from the time tb at which the cranking speed begins to exceed 400 rpm. Thereafter, at the time tc when the VTC phase is advanced up to a predetermined phase, fuel injection starts. The phase-advancing operation of the VTC phase (intake valve open timing IVO as well as intake valve closure timing IVC) is continued until the time tc1, thus resulting in a gradual increase in the quantity of air charged into the cylinder. As a consequence, as can be appreciated from a moderate engine speed rise from the time tc in FIG. 17, it is possible to realize a gradual rise in engine power output or engine torque. As a matter of course, simultaneously with a change in the VTC phase (timing changes for IVO and IVC), the fuel injection amount and injection timing are both controlled properly by means of ECU 6. At the time td that engine temperature Te exceeds the predetermined temperature value, for example 60° C., and engine warm-up has been completed and combustion has been stabilized, the processor of ECU 6 inhibits the VTC phase from being held at the phase-advance state, substantially corresponding to BDC. From immediately after the time td, the VTC phase is controlled in the phase-retard direction (that is, toward a phase suited to the normal engine operation of engine 1), thus ensuring a drop in the effective compression ratio, in other words, improved fuel economy.

On hybrid vehicles employing an automatic engine stop-restart system capable of temporarily automatically stopping an internal combustion engine under a specified condition where a selector lever of an automatic transmission is kept in its neutral position, the vehicle speed is zero, the engine speed is an idle speed, and the brake pedal is depressed, and automatically restarting the engine from the vehicle standstill state, the engine stop and restart operation is frequently executed even after completion of engine warm-up. In the case of engine restart operation, engine 1 has already been warmed up and thus engine 1 is in the stable combustion state without executing phase-advance control for the IVC phase. Therefore, it is possible to omit the phase-advancing process of the VTC phase to a phase corresponding to a phase-advance state (=BDC) from the time tb in FIG. 17. As can be seen from the B characteristic curve indicated by the broken line in FIG. 17, with the lapse of time, the VTC phase is gradually shifted or controlled from the phase-retard state (considerably retarded from BDC and corresponding to the low effective compression ratio) to the phase suited to the normal engine operation of engine 1 without the phase-advancing process to the phase-advance state (=BDC). As a result of this, it is possible to prevent uncomfortable noise and vibrations of the vehicle, occurring owing to a rapid engine-torque rise at the beginning of fuel injection during engine startup under the in-cylinder pressure substantially identical to atmospheric pressure. Additionally, by omitting or eliminating the phase-advancing process to the phase-advance state (=BDC), it is possible to effectively reduce an electric power consumption rate of the engine starter or the motor generator.

The effective compression ratio can be controlled by means of either one of the VTC mechanism, the VVL mechanism, and the VEL mechanism. FIG. 18 shows the intake valve lift and event characteristic, which is obtained by the continuously variable valve event and lift (VEL) control mechanism, capable of continuously varying both of valve lift and event from a short event (small working angle) and low valve lift characteristic to a long event (large working angle) and high valve lift characteristic. As can be seen from the characteristic curves of FIG. 18 attained by the intake-valve VEL system, during a cranking period, the intake valve lift is set to the maximum lift state and thus intake valve closure timing IVC is phase-retarded to reduce the work of compression. As soon as the cranking speed exceeds the predetermined speed value such as 400 rpm (see step 368 of FIG. 11), the intake valve lift is set to the minimum lift state and thus intake valve closure timing IVC is phase-advanced to its maximum phase-advance timing to increase certainty in stable combustion. Thereafter, at a point of time at which engine warm-up has been completed and combustion has been stabilized, the intake valve lift is set to an intermediate lift value between the maximum lift value and the minimum lift value, so as to properly retard intake valve closure timing IVC from BDC, and whereby a mechanical friction loss is reduced and fuel economy is improved.

On compression ignition engines, glow plug (a small electric heater) 8 shown in FIG. 1 is located inside the engine cylinder or an electric heater is often provided in the induction system, for preheating the air or promotion of vaporization of the fuel, thereby assisting spontaneous ignition and promoting combustion during an engine starting period, and consequently enhancing the engine startability. Electric power consumed by the electric heater or glow plug 8 is great (e.g., several amperes). When supplying electric power to the electric heater or glow plug 8 during cranking, there is an increased tendency for the cranking speed to be fallen, thereby deteriorating the engine startability. As a countermeasure for a fall in cranking speed, occurring owing to electric power consumed by the electric heater or glow plug 8, the glow-plug/electric-heater control routine shown in FIG. 19 is executed.

Returning to FIG. 19, at step 391, an ignition switch (an engine key switch) is turned ON. At step 392, electric power supply to glow plug 8 (electric heater) is enabled, and thus electric power is supplied to glow plug 8 (electric heater) to energize it. Subsequently to step 392, step 393 occurs. At step 393, starter 7 becomes energized (ON). At the same time, at step 394, electric power supply to glow plug 8 (electric heater) is shut off (disabled) or reduced to a low level. At step 395 after step 394, a check is made to determine whether the latest up-to-date informational data of engine speed Ne, determined based on the sensor signal from crank angle sensor 5, exceeds a first predetermined speed value such as 400 rpm. When the answer to step 395 is affirmative (YES), that is, when the current engine speed exceeds the first predetermined speed value (e.g., 400 rpm), the routine proceeds from step 395 to step 396. Conversely when the answer to step 395 is negative (NO), step 395 is repeatedly executed, until the current engine speed exceeds the first predetermined speed value owing to a rise in cranking speed. Under the condition of cranking speed above 400 rpm, through step 396 electric power is supplied again to glow plug 8 (electric heater) to energize it, and simultaneously fuel injection starts. At this time (at step 396), in s similar manner to step 369 of FIG. 11, phase-advance control for variable valve actuation mechanism 13 (the VTC mechanism) is executed. Thereafter, at step 397, a check is made to determine whether the latest up-to-date informational data of engine speed Ne exceeds a second predetermined speed value such as 600 revolutions per minute. When the answer to step 397 is affirmative (YES), that is, when the current engine speed exceeds the second predetermined speed value (e.g., 600 rpm), the routine proceeds from step 397 to step 398. Conversely when the answer to step 397 is negative (NO), step 397 is repeatedly executed, until the current engine speed exceeds the second predetermined speed value (e.g., 600 rpm). Immediately when the current engine speed exceeds the second predetermined speed value (e.g., 600 rpm), electric power supply to glow plug 8 (electric heater) is shut off (disabled) through step 398.

By way of execution of the glow-plug/electric-heater control routine shown in FIG. 19, electric power supply to glow plug 8 (electric heater) can be temporarily shut off (disabled) or reduced to a low level, until the cranking speed reaches the first predetermined speed value such as 400 rpm. This increases certainty in sufficient electric power supply to starter 7, thus more certainly enhancing the engine startability.

As will be appreciated from the above, according to the compression ignition engine of the embodiment, employing a variable valve operating system being responsive to a control signal from an electronic control unit for variably adjusting or bringing an intake valve characteristic including at least one of an intake valve lift and an intake valve closure timing IVC closer to a desired value (a desired valve characteristic value determined based on engine operating conditions) via an actuator (electrically-controlled actuator means), during a cranking period of cold starting operation with a starter energized (ON), an effective compression ratio of an engine is temporarily decreased or lowered by controlling the intake valve characteristic. At a point of time when a predetermined cranking speed threshold value (e.g., 400 revolutions per minute) has been reached owing to a cranking speed rise, the effective compression ratio is increased or risen by controlling the intake valve characteristic. After combustion of the engine has been stabilized, the intake valve characteristic is brought closer to the desired value (the desired valve characteristic value) determined based on the engine operating conditions by way of closed-loop control. Thus, it is possible to reconcile the enhanced engine startability during cranking and cold starting operation and improved fuel economy during normal engine operation (after engine warm-up). Suppose that the compression ignition engine of the embodiment, capable of properly controlling the effective compression ratio by varying the intake valve characteristic depending on engine operating conditions, such as during a cranking period of cold starting operation, during an engine warm-up period, and after engine warm-up, is combined with an engine starter of a low torque capacity (or a motor generator of a low torque capacity) and a fixed compression-ratio compression-ignition internal combustion engine of a low geometrical compression ratio. This contributes to the reduced engine gross weight. Thus, the compression ignition engine of the embodiment is suitable for the engine for hybrid vehicles.

Furthermore, according to the compression ignition engine of the embodiment, the variable valve operating system is comprised of a variable valve actuation mechanism capable of varying the intake valve characteristic including at least one of the intake valve lift and intake valve closure timing IVC, before the start of cranking operation or simultaneously with cranking operation, and an engine sensor (concretely, a camshaft sensor) that is able to detect information regarding an intake valve operating state (i.e., the actual intake valve lift and the actual intake valve closure timing) from a substantially zero engine speed value. Thus, even when a temporary drop in battery voltage is occurring during operation of the engine starter, it is possible to satisfactorily adjust (phase-advance or phase-retard) or bring the actual intake valve characteristic, in particular, intake valve closure timing IVC, closer to the desired value, according to various situations, that is, during cranking and starting operation and after warm-up (in a stable combusting state).

Additionally, according to the compression ignition engine of the embodiment, by means of the electronic control unit, at least one of a fuel injection amount and a fuel injection timing, both determined based on engine speed and engine load (e.g., an accelerator-pedal depression amount), can be compensated for based on at least one of information regarding a quantity of air charged into an engine cylinder and information regarding an intake valve operating state (i.e., the actual intake valve lift and the actual intake valve closure timing). Thus, it is possible to compensate for at least one of the fuel injection amount and fuel injection timing in real time responsively to a change in the intake valve operating state, and whereby the generation of soot and unstable combustion can be prevented beforehand.

Additionally, according to the compression ignition engine of the embodiment, when restarting the engine by either one of a starter and a motor generator, an intake valve operating state including at least an actual intake valve closure timing, is gradually shifted or controlled from a phase-retard state to a normal intake valve operating state with the lapse of time. This results in a compression pressure fall of the engine during cranking. Thus, it is possible to reduce the electric power consumption during the cranking period of engine restarting operation, and also to avoid a rapid engine torque rise and uncomfortable noise and vibrations of the vehicle during the restarting operation.

Preferably, during an engine stopping period, the electronic control unit generates a control command signal to the electrically-controlled actuator means for controlling at least the intake valve closure timing IVC to a desired standby timing spaced apart from BDC, and thereafter generates an engine stop signal. During the next starting operation, (i) a check for a current phase of the variable valve actuation mechanism and (ii) phase-retard control of the variable valve actuation mechanism can be eliminated, thereby shortening the engine starting time.

More preferably, during the cranking period of cold starting operation, the electronic control unit operates to temporarily shut off (disable) or reduce electric power supply to either one of glow plug 8 and an electric heater. The temporary shut-off/reduction operation of electric power supply to glow plug 8 or the electric heater, contributes to the increased certainty in sufficient electric power supply to the starter, thereby ensuring the enhanced engine startability.

Moreover, in the case of a starting-period decompression device is combined with the variable valve actuation mechanism, during a cranking period of cold starting operation an exhaust valve is maintained in a constantly-opened valve operating state (i.e., in a decompression mode) to decrease an effective compression ratio by way of decompression for in-cylinder pressure during the cranking period for a smooth cranking speed rise. At a point of time when a predetermined cranking speed threshold value (e.g., 400 revolutions per minute) has been reached owing to the smooth cranking speed rise, the decompression mode is inhibited and the exhaust valve is returned to a normal valve operating state. Additionally, substantially at the point of time when the predetermined cranking speed threshold value (400 rpm) has been reached owing to the smooth cranking speed rise, the effective compression ratio is increased or risen by controlling an intake valve characteristic including at least one of an intake valve lift and intake valve closure timing IVC, for enhancing the self-ignitability of fuel, which is injected after the predetermined cranking speed threshold value (e.g., 400 rpm) has been reached. After combustion of the engine has been stabilized, the intake valve characteristic is brought closer to a desired value (a desired valve characteristic value) determined based on engine operating conditions by way of closed-loop control. By such a combination of the decompression device and the variable valve actuation mechanism, the control system for the variable valve actuation mechanism employed in the variable valve operating system can be simplified, thus ensuring the reduced control system cost. Additionally, by way of an adequate decompressing function of the decompression device, the work of compression can be remarkably reduced, thus enabling a smooth cranking speed rise, that is, a shortened engine starting time.

The entire contents of Japanese Patent Application No. 2005-166538 (filed Jun. 7, 2005) are incorporated herein by reference.

While the foregoing is a description of the preferred embodiments carried out the invention, it will be understood that the invention is not limited to the particular embodiments shown and described herein, but that various changes and modifications may be made without departing from the scope or spirit of this invention as defined by the following claims. 

1. A compression ignition-engine comprising: sensors that detect engine operating conditions; a variable valve operating system comprising at least a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and actuated by an actuator; and a control unit configured to be electrically connected to the sensors and the actuator for controlling the variable valve actuation mechanism via the actuator to bring the intake valve characteristic closer to a desired value determined based on the engine operating conditions detected by the sensors; said control unit comprising a processor programmed to perform the following, (a) temporarily lowering an effective compression ratio of the engine by controlling the intake valve characteristic during a cranking period of cold starting operation; (b) rising the effective compression ratio by controlling the intake valve characteristic at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise; and (c) bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.
 2. The compression ignition engine as claimed in claim 1, wherein: the variable valve actuation mechanism of the variable valve operating system is able to vary the intake valve characteristic, before cranking operation of the engine is started or simultaneously with the engine cranking operation; and the variable valve operating system further comprises a sensor that is able to detect information regarding an operating state of the intake valve from a substantially zero engine speed value.
 3. The compression ignition engine as claimed in claim 1, further comprising: a sensor that is able to detect information regarding an operating state of the intake valve from a substantially zero engine speed value; and a sensor that detects information regarding a quantity of air charged into an engine cylinder, wherein the processor is further programmed for: (d) compensating for, based on at least one of the information regarding the intake valve operating state and the information regarding the quantity of air charged into the cylinder, at least one of a fuel injection amount and a fuel injection timing, both determined based on the engine operating conditions including engine speed and engine load.
 4. The compression ignition engine as claimed in claim 1, further comprising: a sensor that is able to detect information regarding an operating state of the intake valve from a substantially zero engine speed value, wherein the processor is further programmed for: (e) gradually controlling the intake valve operating state including at least an actual intake valve closure timing, to a normal intake valve operating state, when restarting the engine by either one of a starter and a motor generator.
 5. The compression ignition engine as claimed in claim 1, wherein the processor is further programmed for: (f) generating a control command signal to the actuator for controlling at least the intake valve closure timing to a desired standby timing spaced apart from a piston bottom dead center position during a stopping period of the engine; and (g) generating an engine stop signal after generating the control command signal.
 6. The compression ignition engine as claimed in claim 1, wherein the processor is further programmed for: (h) temporarily shutting off or reducing electric power supply to either one of a glow plug and an electric heater, during the cranking period of cold starting operation.
 7. A compression ignition engine (1) comprising: sensors that detect engine operating conditions; a variable valve operating system comprising at least a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and actuated by an actuator; a decompression device provided to operate an exhaust valve in a decompression mode corresponding to a constantly-opened valve operating state during a cranking period of cold starting operation; and a control unit configured to be electrically connected to the sensors and the actuator for controlling the variable valve actuation mechanism via the actuator to bring the intake valve characteristic closer to a desired value determined based on the engine operating conditions detected by the sensors; said control unit also configured to be electrically connected to the decompression device for switching the exhaust valve to the decompression mode during the cranking period; and said control unit comprising a processor programmed to perform the following, (a) temporarily lowering an effective compression ratio of the engine by maintaining the exhaust valve in the decompression mode corresponding to the constantly-opened valve operating state during the cranking period; (b) inhibiting the decompression mode and returning the exhaust valve to a normal operating state at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise; (c) rising the effective compression ratio by controlling the intake valve characteristic substantially at the point of time when the predetermined cranking speed threshold value has been reached owing to the cranking speed rise; and (d) bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.
 8. A compression ignition engine comprising: sensor means for detecting engine operating conditions; a variable valve operating system comprising at least a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and actuated by an actuator; and a control unit configured to be electrically connected to the sensor means and the actuator for controlling the variable valve actuation mechanism via the actuator to bring the intake valve characteristic closer to a desired value determined based on the engine operating conditions detected by the sensor means; said control unit comprising (a) means for temporarily lowering an effective compression ratio of the engine by controlling the intake valve characteristic during a cranking period of cold starting operation; (b) means for rising the effective compression ratio by controlling the intake valve characteristic at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise; and (c) means for bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.
 9. A compression ignition engine comprising: sensor means for detecting engine operating conditions; a variable valve operating system comprising at least a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and actuated by an actuator; a decompression device provided to operate an exhaust valve in a decompression mode corresponding to a constantly-opened valve operating state during a cranking period of cold starting operation; and a control unit configured to be electrically connected to the sensor means and the actuator for controlling the variable valve actuation mechanism via the actuator to bring the intake valve characteristic closer to a desired value determined based on the engine operating conditions detected by the sensor means; said control unit also configured to be electrically connected to the decompression device for switching the exhaust valve to the decompression mode during the cranking period; and said control unit comprising (a) means for temporarily lowering an effective compression ratio of the engine by maintaining the exhaust valve in the decompression mode corresponding to the constantly-opened valve operating state during the cranking period; (b) means for inhibiting the decompression mode and returning the exhaust valve to a normal operating state at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise; (c) means for rising the effective compression ratio by controlling the intake valve characteristic substantially at the point of time when the predetermined cranking speed threshold value has been reached owing to the cranking speed rise; and (d) means for bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.
 10. A method for controlling a compression ignition engine employing a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve, the method comprising: (a) temporarily lowering an effective compression ratio of the engine by controlling the intake valve characteristic during a cranking period of cold starting operation; (b) rising the effective compression ratio by controlling the intake valve characteristic at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise; and (c) bringing the intake valve characteristic closer to a desired value determined based on engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized.
 11. A method for controlling a compression ignition engine employing a variable valve actuation mechanism variably adjusting an intake valve characteristic including at least one of a valve lift of an intake valve and a valve closure timing of the intake valve and a decompression device provided to operate an exhaust valve in a decompression mode corresponding to a constantly-opened valve operating state during a cranking period of cold starting operation, the method comprising: (a) temporarily lowering an effective compression ratio of the engine by maintaining the exhaust valve in the decompression mode corresponding to the constantly-opened valve operating state during the cranking period; (b) inhibiting the decompression mode and returning the exhaust valve to a normal operating state at a point of time when a predetermined cranking speed threshold value has been reached owing to a cranking speed rise; (c) rising the effective compression ratio by controlling the intake valve characteristic substantially at the point of time when the predetermined cranking speed threshold value has been reached owing to the cranking speed rise; and (d) bringing the intake valve characteristic closer to the desired value determined based on the engine operating conditions by way of closed-loop control, after combustion of the engine has been stabilized. 